Opposed free piston engine having start, stop, and restart control means

ABSTRACT

In an opposed free piston engine, each of the engine pistons is formed with a circumferential groove which in one position of the piston registers with a retractable plunger which is radially movable from a retracted position to a piston engaging position, the plunger being normally retained by a spring both in the retracted position and the piston engaging position and being urged into its piston engaging position by hydraulic means.

United States Patent 1191 Fitzgerald [4 1 Sept. 30, 1975 1 OPPOSED FREE PISTON ENGINE HAVING START, STOP, AND RESTART CONTROL MEANS [76] Inventor: William Maurice Bard Fitzgerald,

R.R. No, l', Claremount, Ontario, Canada 22 Filed: Sept. 9, 1974 211 Appl. No.: 504,729

Related U.S. Application Data [62] Division of Ser. Nov 305,453, Nov. 10, 1972, Pat. No.

[52] U.S. Cl 60/595; 91/41 [51] Int. Cl. F021) 71/04 [58] Field of Search 60/595, 596, 416, 371; 91/41, 43, 45

[56] References Cited UNITED STATES PATENTS Vickers 91/45 Huber 60/596 x 2,795927 6/1957 Huber.. 60/596 2,839,911 6/1958 Smith 60/596 3,365,879 l/l968 Panhard 60/595 Primary Examiner-Edgar W. Geoghegan [57] ABSTRACT In an opposed free piston engine, each of the engine pistons is formed with a circumferential groove which in one position of the piston registers with a retractable plunger which is radially movable from a retracted position to a piston engaging position, the plunger being normally retained by a spring both in the retracted position and the piston engaging position and being urged into its piston engaging position by hydraulic means.

9 Claims, 44 Drawing Figures US. Patent Sept. 30,1975 Sheet10f14 3,908,379

U.S. Pa tent Sept. 30,1975 Sheet2of14 3,908,379

Fir-1n U.S. Patent Sept. 30,1975 Sheet3of 14 3,908,379

US. Patent Sept. 30,1975 Sheet4of 14 3,908,379

U.S. Patent Sept. 30,1975 SheetSof 14 3,908,379

237 2'40 287 104L 104R F l 6. 2O

205 F I G. 18 299 US. Patent Sept. 30,1975 Sheet7of 14 3,908,379

204 mp 294 44 293 372 29s FIG. 30

US. Patent Sept. 30,1975 Sheet8of 14 3,908,379

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US. Patent Sept. 30,1975 Sheet 10 of 14 3,908,379

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U.S. Patent Sept. 30,1975 Sheet110f14 3,908,379

FIG. 35

U.S. Patent Sept. 30,1975 Sheet 12 of 14 3,908,379

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US. Patent Sept. 30,1975 Sheet 13 of 14 3,908,379

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OPPOSED FREE PISTON ENGINE HAVING START, STOP, AND RESTART CONTROL MEANS CROSS REFERENCE TO RELATED APPLICATION The present application is a divisional of my copending application Ser. No. 305,453, filed Nov. 10, 1972 now Pat. No. 3,841,797, issued Oct. l5, 1974 relating to Power Units.

BACKGROUND OF THE INVENTION This invention relates to free piston engines of the type having a pair of opposed free pistons working in a common cylinder. Such a free piston engine is described in my copending application Ser. No. 305,453 identified above.

Hitherto, free piston engines of this type have been proposed, but the proposals have not included any satisfactory means for starting the engine, and no provision has been made for stopping the engine in such a manner as to ensure that the pistons would stop in a suitable position for restarting of the engine.

The object of the present invention is to provide a free piston engine in which such provision for starting, stopping, and restarting are provided.

SUMMARY OF THE INVENTION A free piston engine according to the invention is characterized in that each of the engine pistons is formed with a circumferential groove which in one position of the piston registers with a retractable plunger which is radially movable from a retracted position to a piston engaging position, the plunger being normally retained by a spring in the retracted position and being urged into its piston engaging position by hydraulic means.

One embodiment of the invention, as applied to a power unit and transmission system for a wheeled vehicle, will now be described by way of example with reference to the accompanying drawings, in which:

FIG. 1 is a longitudinal section taken through the axis of the power unit;

FIG. 2 is a fragmentary top plan view of the power unit;

FIGS. 9, 10, 11, 12 and 13 show details of a plate valve assembly shown in FIGS. 1, 6 and 8;

FIG. 14 is an underneath plan view of the detail shown in section in FIG. 6, without the inlet valve assembly;

FIG. 15 is a diagrammatic drawing showing leakage control of oil from the pump pistons of the power unit; FIG. 16 shows a section on line 16-16 in FIG. 1; FIG. 17 shows a section on line 17-17 in FIG. 2;

FIG. 18 shows a central vertical section through the fuel injector of the unit, the section being on line 18-18 of FIG. 19;

FIG. 19 is a partly sectioned side elevation of the fuel injector;

FIG. 20 is a section on line 20-20 in FIG. 18;

FIG. 21 is an unsectioned end view of FIG. 18;

FIG. 22 shows the outside of a spool valve element;

FIG. 23 is a section on line 23-23 in FIG. 18;

FIG. 24 is a section on line 24-24 in FIG. 18;

FIG. 25 is a section on line 25-25 in FIG. 18;

FIG. 26 is a sectional view of a valve connector adapted to be used with the fuel injector;

FIG. 27 is a part-sectional plan view of a control gear for starting and stopping the power unit;

FIG. 28 is a part elevation on line 28-28 in FIG. 27;

FIG. 29 is an unbroken plan view of the control gear;

FIG. 30 is an end elevation of the control gear viewed from the left in FIG. 27 with certain parts removed;

FIG. 31 is a sectional plan view on line 31-31 in FIG. 27;

FIG. 32 is a fragmentary view in the direction of arrow 32 in FIG. 27 with certain parts removed;

FIG. 33 is a part elevation showing the end view of a solenoid;

FIG. 34 is a partly broken away side elevation of a reversible hydraulic motor adapted for use with the power unit;

FIG. 35 is a section on line 35-35 in FIG. 34;

FIGS. 36 and 37 illustrate details of a planet gear bearing from opposite sides thereof;

FIG. 38 is a section on line 38-38 in FIG. 37;

FIG. 39 is a section on line 39-39 in FIG. 34 with certain parts removed, showing a planet gear without details of its teeth;

FIG. 40 is a schematic view of a gear wheel of the motor; and

FIGS. 41a and 41b are a schematic overall representation of the complete power unit and ancillary equipment;

FIG. 42 is a section on line 42-42 in FIG. 41;

FIG. 43 is a section on line 43-43 in FIG. 41.

THE POWER UNIT General The power unit comprises an internal combustion engine having a pair of opposed free pistons, a pair of pump units the pistons of which coact with the engine pistons, a pair of constant displacement hydraulic accumulators into which pressurized hydraulic fluid is pumped in accordance with the expansion strokes of the engine pistons, inlet ports and exhaust ports under the control of the pistons for admitting combustion air to, and exhausting combustion gases from, the engine, and valve operated fuel injection means actuated in accordance with the cyclical movements of the pistons to control the injection of fuel into the engine. The power output from the engine is a flow of pressurized hydraulic fluid, which in the present example is delivered from a pair of smoothing accumulators and used to drive hydraulic motors.

ARRANGEMENT AND MECHANICAL CONSTRUCTION The mechanical construction of the power unit itself, and certain details of such construction, are illustrated in FIGS. 1 to 17, of which FIG. 1 best illustrates the general arrangement of the unit. Reference will now be made to these figures in particular.

At the heart of the power unit is a compressionignition engine comprising a single, water-cooled cylinder 101 having a ring of air inlet ports 102 and a ring of exhaust ports 103, and a pair of opposed free pistons 104L and 104R of equal mass. The pistons 104L and 104R are free to reciprocate within the cylinder 101, and the overall design includes means to ensure that the pistons always move simultaneously in opposite directions and are also disposed symmetrically on opposite sides of a central position denoted by line 4-4 in FIG. 1. A fuel injector 200 is bolted to the cylinder 101 at the central position, so that its nozzle 201 is positioned to inject fuel into the space between the opposed pistons at appropriate times, as will be described hereinafter.

Each end of the cylinder 101 is bolted to a respective one of two similar hydraulic accumulator-pump assemblies 105, 106. The assembly 105 (which will be described in detail, the assembly 106 being identical in construction) comprises a pump unit 107, a first, constant displacement hydraulic accumulator 108, and a second, high pressure or smoothing hydraulic accumulator 109, the assembly having a casing structure including a bulkhead 110 which is bolted to the cylinder end by bolts 111.

The pump unit 107 provides an internal oil-filled space or pump chamber 112, and houses a composite cylindrical or pump piston 113 which is a reasonably leak-free sliding fit in the bulkhead 110. The combined effects of the momentum of the pump piston 113, and the pressure in the pump chamber 112, ensure that the pump piston 113 is always pressed against the engine piston 104L.

A groove 114 in the bulkhead allows oil leaking along the outer wall of the pump piston 113 to pass into a pipe 177 (FIG. 5) which conveys it to a float chamber 178. Within the float chamber 178 is a float 179, which on rising uncovers a drain hole 180 leading back to a vented reservoir 514, (FIG. 41) via a pipe X. A tube, 181 leading via a restrictor 182 to a hole 183 in the bulkhead 1 gives access to the air compression space 117. A valve 184 may be used to open or close access to the air compression space so that air may be extracted therefrom in order to form a vacuum with which to suck the pistons 104L and 104R into their starting position. The float 179 and drain hole 180 are preferably so dimensioned that the float will rise before its total submersion under any conditions of average pressure that may exist in float chamber 178. In any case after the engine is brought to rest, pressure in the float chamber 178 will fall so that oil leaking along the outer wall of the pump piston 113 will flow into it, and when the level is sufficiently high the float will rise and allow the leaking oil to flow down into a vented reservoir 514 (FIG. 41).

The pump piston 113 is in the form of a hollow ram which defines an internal oil space and contains a heavy plunger 115 which is free to move back to a retaining screw 118, and forward to cover an oil flow restrictor 1 16 at the inner end of the ram; it will move in this manner under the impetus of its own inertia, pressing against the oil flow restrictor 116 at the inner end when the piston 104L is accelerating during the first part of its outward stroke and decelerating during the last part of its re-compression stroke. The ram will be pressed against the retaining screw 118 during the last part of its expansion stroke and the first part of its recompression stroke. The restrictor 116 allows a controlled quantity of oil to pass into a hole shown by dotted lines 164 and on through a number of holes 165 into a groove 166 round the piston 104L for cylinder wall lubrication where the piston slides. A spring 118A contained within the retaining screw 118 urges the solid rod 115 inwards to close the restrictor 116 when the engine is at rest so that oil cannot escape through the restrictor at this time.

The first, constant displacement accumulator 108 comprises a domed casing providing a stepped cylindrical internal surface 120. The domed casing houses a downwardly projecting cylindrical sleeve 121 in which a piston 122 is free to move axially up or down. The lower portion of said stepped cylindrical surface constitutes a cylinder communicating with the pump chamber 112 and locating a leak-free piston 123 which is free to move axially up or down. The pistons 122 and 123 define within the first accumulator a space 124 of variable volume which contains nitrogen or other gas under pressure. A second, oil-filled space 125 is situated above piston 122, to or from which oil may be admitted or withdrawn via a port 126 in the dome 127 of the casing. The piston 123 is formed with a flange 128 which is adapted to come to rest against a step 129 of said stepped cylindrical surface 120 and to abut against the lower end of the sleeve 121, for limiting the downward and upward movements of the piston 123. Thus the piston 123 is constrained to move between lower and upper limit positions which determine the minimum and maximum charge levels of the accumulator, respectively. A sump 130 is formed by the piston 123, in which any oil that may leak into the space 124 will collect and from which it may be withdrawn via a duct 131. This duct also serves for recharging the gas space 124 and to adjust its pressure; a suitable valve would normally be fitted into the duct 131.

It is necessary to ensure rapid establishment of inlet oil flow into the pump chamber 112 once the chamber pressure falls as the result of the flange 128 of accumulator piston 123 coming to rest against the step 129, while the pump piston 113 still continues its outward stroke. For this purpose, an inlet oil assembly 132 is provided to control the flow of oil through a springloaded plate valve 133. In this assembly: (1) the mass of the moving element of the valve 133 is kept reason ably low; (2) a diaphragm 132D, backed by a suitable gas such as nitrogen, contained in a space 1326, keeps to a minimum the mass of oil that must be accelerated on each cycle; (3) the cross-sectional area of the oil, perpendicular to its direction of flow, is large so as to keep the oil velocity low; (4) the inlet oil in space 132E is raised to a fairly high pressure, which for example might in a particular instance be pounds per square inch. When the pump piston 113 moves inwards into the pump chamber 112, the first accumulator is first charged to its full capacity and then oil is forced via a second automatic spring-loaded plate valve 134 into the second hydraulic accumulator 109. The details of construction of the second automatic plate valve 134, which is essentially a high speed one-way valve, are illustrated in FIGS. 9 to 13. The first automatic plate valve may be similarly constructed. As illustrated in FIGS. 9 to 13, the valve comprises essentially a stationary valve element in the form of a grid, and a thin plate which is formed as a complementary grid, the thin plate being urged into contact with the stationary valve element by an array of compression springs. When the valve is closed, the grid elements of the thin plate close off the spaces between the elements of the stationary valve element; when the thin plate is displaced by a small amount, however, these spaces are opened simultaneously. Thus the valve opens substantially to its maximum extent with a minimal displacement of the movable valve element.

The second hydraulic accumulator 109, best shown in FIG. 6, is connected to the oil delivery opening of the pump chamber 112, which opening is controlled by the automatic one-way valve 134. The accumulator 109 comprises a domed casing housing a cylindrical sleeve 135 within which a piston 136 is free to slide axially up or down. The piston 136 defines within the sleeve 135 a space 136, which is filled with nitrogen or other suitable gas under pressure. A vent 138 (FIG. 17) for filling the space 137 leads into the clearance space that exists between the cylindrical sleeve 135 and the casing 109. When the pump piston 113 moves inwards it first charges the constant displacement recompression accumulator 108 by forming the piston 123 up until the latter is brought to rest against the lower end of the sleeve 121, and oil then passes from the chamber 112 via the one-way valve 134 into an oil space 139 of the second accumulator 109, which has an outlet port 140 from which the pressurized oil is supplied to the hydraulic load circuit.

Communicating with the air compression space 117 behind each of the engine pistons 104L and 104R is a thin spring-loaded air inlet valve 141 fitted in an entrance 142, and a thin spring-loaded air delivery valve 143 fitted in an outlet 144; these valves admit air into the compression spaces 117 on the compression strokes of the pistons, and permit egress of air from the compression spaces on the expansion strokes of the pistons, respectively. The entrances 142 may be connected by flexible metal tubing to an air inlet filter, atmospheric air being filtered and admitted through the valves. In the preferred embodiment illustrated in the drawings, however, the entrances 142 are connected by a duct 145 to the outlet of an air compressor 146. The duct 145 may also include an air cooler. The compressed air from the outlets 144, after cooling of necessary, is conveyed via ducts 147 (shown broken away in FIG. 1) to an inlet 148 communicating via an air inlet manifold 149 with the inlet ports 102 of the engine cylinder (see FIG. 3). The inlet ports 102 and the ends of the air inlet manifold 149 are preferably shaped so as to induce a swirling motion of the incoming air, for example as indicated by the arrows of FIG. 3. In the present embodiment the depth of the channel perpendicular to the cross sectional plane of FIG. 3, tapers from the inlet 148 to the ends of the manifold so as to promote approximately the same air velocity throughout the manifold. This produces two elongated air flow vortices in the engine cylinder, as indicated in FIG. 3. Fuel is injected from the injection nozzle 201, as indicated in FIG. 4, at about maximum compression.

Additional openings 150 may be provided in the wall of the engine cylinder 101. These openings, one of which is presently shown closed by a cover 151, may be used to apply compressed air for moving the pistons apart if necessary, (when the engine is inoperative,) or to connect a pressure gauge for research or experimental purposes, or to provide an alternative means for fuel injection or fuel injection timing, or to admit air when the pistons are being set in a starting position, as will be explained hereinafter.

The engine exhaust system comprises a casing 152 providing a storage space 153, which communicates with the exhaust ports 103 via a manifold or passage 154 in the engine cylinder. A casing 155 bolted to the bottom end of the casing 152 provides an internal space 156 which communicates with the storage space 153 by way of ports 157. An intake tube 158 connected to the upper end of the casing projects upwards in alignment with the passage 154, the latter being spaced from the end of the intake tube. The casing 155 also provides an internal cylinder portion containing a spool valve 160, which is biassed upwardly by a spring shown diagrammatically at 159. In operation of the engine, when the exhaust ports 103 are uncovered by the piston 104R towards the end of a compression stroke, the products of combustion enter the storage space 153 and impinge upon the intake tube 158. If the kinetic energy of the exhaust gases is relatively low, the spool valve 160 remains in its upper position and the exhaust gases pass to a silencer (not shown) via ducting 161. However, if the kinetic energy of the exhaust gases is sufficiently high, the spool valve 160 is displaced downwards to cover the ports 157; in this case the gases in the storage space 153, being of increased pressure, pass through a duct 162 to an exhaust turbine, the latter being combined with the air compressor 146. The spent gases are finally exhausted via a pipe 163 and a silencer (not shown).

Each of the engine pistons is formed with a circumferential annular groove 169 having bevelled sides, into which a plunger 170 having a correspondingly bevelled end may be formed, in order to lock the pistons against movement when the engine is not running. In FIG. 1 the plungers 170 are shown neither fully in nor fully out, but are shown for illustration in an intermediate position. Each plunger 170 is normally held out of the respective annular groove 169 when the engine is running, by a U spring 171 which engages a grooved portion 172 of the plunger. When required, the plungers 170 are pressed into their operative, piston-holding positions by actuating pistons 173; the latter are slidable in cylinders 174 and are actuated by hydraulic pressure applied via oil connections 175 when the pistons 104L and 104R are near the ends of their expansion strokes. The plungers 170 hold the pistons 104L and 104R approximately in the position shown in FIG. 1, against the forces exerted by the rams 1 13, the latter being urged by hydraulic pressure from the accumulators 108. The forward speed of each actuating piston 173 is controlled by an orifice in a plate valve 167; the parts are designed to permit comparatively free flow in the reverse direction when the engine is being started. A light spring 168 holds the valve plate normally forward.

The output of the engine is a flow of pressurized liquid. The amplitude and frequency at which the engine pistons 104L and 104R reciprocate are variable, depending upon the power that is being developed. The positions at which the pistons momentarily stop at the end of the compression stroke are largely determined by the initial momentum of the pistons and the pressure of the initial cylinder air charge. The positions at which the pistons momentarily stop at the end of the expansion stroke are determined by the momentum that they have gained from the energy of combustion, the cyclic range of oil pressures, and the rate of flow of the hydraulic liquid. To maintain an approximately constant compression ratio in the engine cylinder 101, more energy will be required at high intake air pressures than at low intake air pressures; however, the same invariable volume of oil, determined by the stroke of each piston 123 of the constant displacement hydraulic accumulators, as it reciprocates between the limits of its movement, will always be set aside for the return strokes of the engine pistons. Therefore the average pressure of this oil must be altered as required in accordance with the pressure of the intake air. It can easily be shown that when the average oil pressure in each of the constant displacement accumulators 108 is low, (to accommodate low intake air pressure,) the average compression speeds of the pistons 104L and 104R will also be low and the time taken to effect the compression strokes will be correspondingly long. Conversely, when the average oil pressure in each of the constant displacement accumulators 108 is high, the speeds of the pistons 104L and 104R will be high, so that the time to effect the compression strokes will be correspondingly short. The same factors apply to the speeds of the pistons on the expansion strokes, so that the time to complete the strokes will be an inverse function of the energy developed. The net result is that the rate of reciprocation of the engine pistons will be low at low power outputs and high at high power outputs.

POWER UNIT OPERATION With the stop" plungers 170 completely retracted, the engine pistons 104L and 104R and the pump pistons 113 are initially in the positions shownin FIG. 1 with the engine pistons moving towards one another; the engine pistons 104L and 104R and the pump pistons 113 together have sufficient momentum to give an air compression ratio of, say :1, or higher. Subsequently to this initial stage the oil inlet plate valves 133 open. The air inlet valves 141 are already open and the air delivery valves 143 are closed, so that air enters the air compression spaces in the engine cylinder lying between the engine pistons and the bulkheads 110. The pump pistons 113 also move out from the pump chambers 112, under the action of their momentum and the pressurized fluid passing through the inlet oil valves 133. At about the position when the momentum of the pump pistons, together with the momentum of the engine pistons 1041. and 104R, is spent in compressing the air charge in the cylinder 101, fuel is injected into the cylinder by the fuel injector 200; the gas charge temperature and the relative pressure then rises and the engine pistons are caused to accelerate away from each other, performing their working or expansion stroke.

The one-way plate valves 133 then close and the pump pistons 113 are forced inwards by pistons 104, first to charge the accumulators 108, which are adjusted so as to yield at less pressure than the pistons 136 of the variable displacement accumulators 109. When the pistons 123 have completed their strokes, which are terminated by the abutment of these pistons against the lower ends of the sleeves 121, the pressure in the pump chambers 112 continues to rise and the one-way plate valves 134 open against the pressure in the oil spaces 139 of the variable displacement accumulators 109. During normal running of the unit, the pistons 136 are at a higher position than that shown in the drawings,

depending upon the pressure required. The surge of oil on each pumping stroke, after first displacing the pistons 123 in the constant displacement accumulators 108, is absorbed in urging the pistons 136 inwards against the pressure of gas in the gas spaces 137, but the oil is continually leaving at a more moderate velocity through the outlet ports 140. When the momentum of the engine pistons 104L and 104R is spent, the pistons stop and the high speed plate valves 134 close. The pistons 123 are subjected to pressure from the gas above them and still maintain a substantial pressure on the oil in the pump chambers 121; this pump chamber pressure acts on the pump pistons 113 and accelerates the engine pistons 104L and 104R back along another compression stroke as already described. During the time in which the pistons 104L and 104R are moving outwards on the expansion stroke, the pressure in the air compression spaces 1 17 is increasing, and when this pressure exceeds the pressure above the air delivery valves 143, the latter open to admit this air charge.

Outlets 176 from the pump chambers 112 are connected to pipes 515L, 515R (FIG. 41), as will be described hereinafter. A peripheral groove 119 is incorporated in the pistonrod bearing of each bulkhead as part of a means to ensure synchronisation of the engine pistons 1041.. and 104R, as will also be explained hereinafter. By these means continuous running of the en gine is achieved.

THE FUEL INJECTOR General The fuel injector 200 of FIG. 1 comprises injection means for injecting predetermined quantities of fuel into the engine cylinder, and actuator means for actuating the injection means in accordance with the air pressure within the engine cylinder so as to ensure that the predetermined quantities of fuel are injected into the cylinder at appropriate times in relation to the combustion cycle of the engine.

The injection means comprises, basically, a fuel injection nozzle, a valve controlled supply chamber located behind the injection nozzle, means for admitting fuel to the fuel supply chamber, and a fuel piston actuated by said actuator means to compress the fuel in the supply chamber and to expel the fuel therefrom to the engine cylinder via the injection nozzle. The actuator means includes a spring-loaded shuttle valve arranged to move towards one or other of two limit positions in accordance with the gas pressure in the engine cylin der, and means for supplying pressurized hydraulic fluid to actuate the fuel piston in accordance with the position of the shuttle valve; the pressurized hydraulic fluid is fed from a chamber housing a free piston which is urged in a direction to expel hydraulic fluid from the chamber, expulsion of fluid from the chamber being controlled by the shuttle valve, which is arranged to cover and uncover a port leading to the chamber.

ARRANGEMENT AND MECHANICAL CONSTRUCTION The fuel injector is illustrated in detail in FIGS. 18 to 26 of the drawings, of which FIG. 18 best shows the interrelationship of its working parts. Reference will now be made to these figures in particular.

In FIG. 18 is shown a portion of the engine cylinder 101, and portions of the engine pistons 104L and 104R which define a combustion space S in the engine cylinder into which fuel is injected by the fuel injector. The fuel injector itself is incorporated in a metal body 202, which is machined to provide a number of internal passages and bores as hereinafter described, and which houses the essential elements of the injection means and the actuator means referred to above.

The metal body 202 is formed with a stepped cylindrical bore 203, at the upper end of which is an assembly consisting of the injection nozzle itself 201, a spring-loaded valve 204 having a valve seat 205, a spacer ring 206, and another spacer ring 205A which may be of relatively soft metal such as mild steel, those parts being clamped and retained in position by an adaptor 207 which is screwed into the threaded upper end of the cylindrical bore 203. The adaptor 207 is located in a passage in the wall of the engine cylinder 101, to which the fuel injector body 202 is suitably connected, a sealing ring 208 being located so as to prevent leakage of gases from. the engine cylinder.

Located within the cylindrical bore 203 is a cylindrical barrel 209, the barrel being a tight leak-free fit within the bore. A piston 210 having trunk extension 211 of slightly reduced diameter is slidably arranged within the barrel 209 to define a space 212, constituting a fuel supply chamber, between the upper end of the trunk extension and the valve 204. A compression spring 213 encircling the trunk extension 211 biases the piston 210 towards its lowermost position. Fuel is admitted to the supply chamber 212 through a port 214 in the barrel 203, the port communicating with a supply inlet via a passage 215.

Also located within the cylindrical bore 203 is a second barrel 216, this also being a tight leak-free fit in the bore. The bottom end of the piston 210 is castellated so that, as it is biassed downwardly by the spring 213, a space 219 remains beneath the piston for the admission of hydraulic fluid.

A piston valve 220, which is a low clearance running fit in the barrel 216, is biassed upwardly by a spring 221, the piston valve having a flange 222 against which the spring bears. Upward and downward movements of the piston valve 220 are limited by engagement of the flange 222 with a step 223 in the barrel 216, and with a sleeve member 224, respectively. The piston valve 220 is provided for the purpose of allowing rapid egress of used oil from the space 219 through an internal passage 225 of the valve. The passage 225 communicates via radial holes 226 with a shallow annular space 227 near the upper end of the valve. When the valve is in its upper position, the annular space 227 communicates with the space 219 and permits oil egress, and when the valve is in its lower position, the annular space 227 is isolated from the space 219. Thus the valve is closed in its lower position and open in its upper position.

The spring 221 is weaker, in terms of simple force,

than the spring 213. However, in relation to the cross sectional areas of the piston valve 220 and piston 210, respectively, against which the springs act, the spring 221 is the stronger of the two. When therefore the piston 210 is moving outwards to expel the spent oil in space 219, the piston valve 220 will be open but when the castellated end of the piston 210 reaches the inwardly extended end of the piston valve 220, the latter will cade and close. Thereafter introduction of high pressure oil into the space 219 will oppose the piston 210 and move it upwards, while simultaneously it will keep the piston valve 220 closed down.

In practice a very small leakage of oil from the space 219 is required when the piston valve 220 is closed, and for this purpose the upper end of the piston valve may contain a small longitudinal channel of appropriate cross section.

Any leakage of oil into the annular space between the piston extension 211 and the inner surface of the barrel 209 can pass out through a small hole 228 and a oneway valve consisting of an O-ring 229 located in an annular groove in the outer surface of the barrel 209.

A screw threaded hole 231 in the metal body 202 communicating with the passage 215 is adapted to receive a connector valve 232 (FIG. 26) whereby fuel is admitted to the passage 215 and thence to the supply chamber 212 via the port 214. The connector valve 232 comprises a valve body having a first stem portion 233, which is adapted to be screwed into the hole 231, a second stern portion to which a fuel supply line may be connected, and a flange 234 which is adapted to bear against the fuel injector body 202. Within the valve body is a valve member 235 biassed towards its closed position by a spring 236.

The fuel injector body 202 is also formed with a second cylindrical bore 237, housing a barrel 238 which is a tight stationary leak-free fit within the bore. Mounted within the barrel 238 is a spring-loaded shuttle valve 239, to which is connected a downwardly extending hollow rod 240. A carrier cup 242 is free to move axially to and fro in a sleeve 248. A compression spring 241 contained within the carrier cup 242 acts upon the rod 240 to urge the shuttle valve 239 to its uppermost position, at which position the valve lies very close to, but is spaced from, the machined outer surface of the engine cylinder 1 01. The upper limiting position of the shuttle valve 239 is determined by the abutment of the carrier cup 242 against a flange 243 of a third barrel 244, the diameter of the flange 243 ensuring a tight leak-free fit in the fuel injector body 202. A central hole 245 in the top of the carrier cup 242 is aligned with the hollow rod 240 for receiving oil which leaks down the rod, this oil passing via a port 246 to the drainage passage 218. Downward movement of the carrier cup 242 is limited by a step 247 on the inner surface of the sleeve 248.

Means are provided to ensure that the port 246 remains in line with the corresponding hole in the fuel injector body 202, so that leakage oil may pass freely to the passage 218. The flange 243 is kept firmly locked against its seat in the body 202 by a screw 251, acting through the sleeve 248. A seal 299 is fitted round the top of the barrel 238 to prevent leakage of cylinder gas.

A passage 249 in the engine cylinder extends from the cylinder space S (FIG. 18) to the space immediately above the shuttle valve 239, so that the latter is exposed to cylinder gas pressure and will be caused to move downwardly in the barrel 238 when the gas pressure exceeds a value determined by the force exerted on the shuttle valve by the spring 241. The force exerted by the spring 241, and hence the value of cylinder pressure at which the shuttle valve is moved downwards, can be pre-adjusted by any suitable means, such as an adjustable screw plug 250, located in the bottom end of the retaining screw 251.

The upper end of the barrel 244 is of reduced diameter so as to be a tight leak-free fit in a recess machined into the lower end of the barrel 238.

The working oil at suitable pressure is supplied to the fuel injector by way of a feed pipe (not shown) connected to an inlet adaptor 253 including a springloaded one-way valve 254. The oil passes via a passage 255 to an annular space 256 surrounding the stem 257 of theyshuttle valve 239, and thence through a passage 

1. A power unit comprising a free piston engine having a pair of opposed free pistons working in a common cylinder, each of the engine pistons being formed with a circumferential groove which in one position of the piston registers with a retractable plunger which is radially movable from a retracted position to a piston-engaging position, the plunger being normally retained by a spring both in the retracted position and the piston-engaging position and being urged into its piston-engaging position by hydraulic means.
 2. A power unit according to claim 1, wherein that the plungers are urged into their engaging positions by respective actuating pistons operated by hydraulic fluid under pressure, the supply of hydraulic fluid to the actuating pistons being controlled by valve gear in response to movements of an operating member between a ''''stop'''' position and a ''''running'''' position.
 3. A power unit according to claim 2, said valve gear comprising a first spool valve movable from a first position to a second position to admit hydraulic fluid to the actuating pistons, a hydraulic ram for moving the first spool valve from its first position to tis second position, and a second spool valve movable from a first position to a second position to admit hydraulic fluid to said hydraulic ram, the second spool valve being normally biassed towards the second position and displaceable therefrom in response to movement of the operating member from said ''''stop'''' position.
 4. A power unit according to claim 3, said valve gear including a slide member engaging the second spool valve, the slide member being urged from a neutral position by the operating member upon movement of the operating member from the ''''stop'''' position to displace the second spool valve from its second position.
 5. A power unit according to claim 4, including a bolt engageable with the first spool valve and operable to prevent movement of the first spool valve to its second position, the bolt being urged into its engaging position in response to fluid pressure and released to permit movement of the first spool valve to its second position when the fluid pressure is a minimum.
 6. A power unit according to claim 5, wherein the bolt is urged to its engaging position by hydraulic pressure controlled by a valve responsive to said fluid pressure.
 7. A power unit according to claim 5 wherein the speed of movement of the first spool valve is controlled by an orifice so as to operate the actuating pistons after a predetermined delay time.
 8. A power unit according to claim 4, wherein the slide member is engaged by a spring-loaded detent which initially resists movement of the slide member from its neutral position.
 9. A power unit according to claim 2, including a pair of spring-loaded bolts engageable with the operating member to prevent movement of the operating member from the ''''stop'''' Position, the spring-loaded bolts being opposed by hydraulic rams responsive to the pressures of accumulators of the power unit and operative to prevent movement of the operating member to the ''''running'''' position when the pressures of the accumulators are less than a predetermined value. 